JP4428442B2 - Spark ignition internal combustion engine - Google Patents

Description

  The present invention relates to a spark ignition internal combustion engine.

Combustion chamber volume variable device is equipped, and when the temperature of the catalyst arranged in the engine exhaust passage is lowered and the activity is likely to be lost, the combustion chamber volume is increased and the actual compression ratio is lowered, thereby causing combustion. A spark ignition type internal combustion engine in which efficiency is deteriorated to increase exhaust gas temperature to increase the temperature of a catalyst is known (see, for example, Patent Document 1).
Japanese Patent Publication No. 4-28893

  However, if the actual compression ratio is lowered in order to increase the temperature of the catalyst, there is a problem that stable combustion cannot be obtained because the ignition and combustion of the fuel deteriorates.

In order to solve the above problem, according to the present invention, a variable compression ratio mechanism capable of changing the mechanical compression ratio, a variable valve timing mechanism capable of controlling the closing timing of the intake valve, and the engine intake passage are disposed. A throttle valve and a catalyst disposed in the engine exhaust passage are provided. When the catalyst is activated, the mechanical compression ratio is increased to the maximum mechanical compression ratio as the engine load is reduced, and the intake valve is closed. In the spark ignition type internal combustion engine that is moved in a direction away from the intake bottom dead center as the engine load becomes low, the valve timing includes a predicting means for predicting the temperature of the catalyst disposed in the engine exhaust passage. engine load in order to reduce the actual expansion ratio while increasing the while being or actual compression ratio holding the actual compression ratio at the same is lowered when the temperature of the catalyst is expected to drop below the activation temperature Correspondingly, the larger the amount of decrease in the mechanical compression ratio, the larger the amount to be moved toward the valve closing timing of the intake valve to the intake bottom dead center, the amount of decrease in the opening degree of the throttle valve is increased.

  The temperature of the catalyst can be increased while ensuring good fuel ignition and combustion.

FIG. 1 shows a side sectional view of a spark ignition type internal combustion engine.
Referring to FIG. 1, 1 is a crankcase, 2 is a cylinder block, 3 is a cylinder head, 4 is a piston, 5 is a combustion chamber, 6 is a spark plug disposed at the center of the top surface of the combustion chamber 5, and 7 is intake air. 8 is an intake port, 9 is an exhaust valve, and 10 is an exhaust port. The intake port 8 is connected to a surge tank 12 via an intake branch pipe 11, and a fuel injection valve 13 for injecting fuel into the corresponding intake port 8 is arranged in each intake branch pipe 11. The fuel injection valve 13 may be arranged in each combustion chamber 5 instead of being attached to each intake branch pipe 11.

  The surge tank 12 is connected to the outlet of the compressor 15a of the exhaust turbocharger 15 via an intake duct 14, and the inlet of the compressor 15a is connected to an air cleaner via an intake air amount detector 16 using, for example, heat rays. A throttle valve 19 driven by an actuator 18 is disposed in the intake duct 14.

  On the other hand, the exhaust port 10 is connected to an inlet of an exhaust turbine 15b of an exhaust turbocharger 15 via an exhaust manifold 20, and an outlet of the exhaust turbine 15b is connected to a catalytic converter 22 containing, for example, a three-way catalyst via an exhaust pipe 21. Is done. An air-fuel ratio sensor 23 is arranged in the exhaust pipe 21, and a temperature sensor 24 for detecting the temperature of the three-way catalyst is arranged downstream of the catalytic converter 22.

  On the other hand, in the embodiment shown in FIG. 1, the piston 4 is positioned at the compression top dead center by changing the relative position of the crankcase 1 and the cylinder block 2 in the cylinder axial direction at the connecting portion between the crankcase 1 and the cylinder block 2. There is provided a variable compression ratio mechanism A capable of changing the volume of the combustion chamber 5 when performing the operation, and the valve closing timing of the intake valve 7 can be controlled and the intake air can be controlled in order to change the actual start timing of the compression action. A variable valve timing mechanism B capable of individually controlling the valve opening timing of the valve 7 is provided.

  The electronic control unit 30 is composed of a digital computer and includes a ROM (Read Only Memory) 32, a RAM (Random Access Memory) 33, a CPU (Microprocessor) 34, an input port 35 and an output port 36 connected to each other by a bidirectional bus 31. It comprises. The output signal of the intake air amount detector 16, the output signal of the air-fuel ratio sensor 23, and the output signal of the temperature sensor 24 are input to the input port 35 via corresponding AD converters 37, respectively. A load sensor 41 that generates an output voltage proportional to the depression amount L of the accelerator pedal 40 is connected to the accelerator pedal 40, and the output voltage of the load sensor 41 is input to the input port 35 via the corresponding AD converter 37. Is done. Further, a crank angle sensor 42 that generates an output pulse every time the crankshaft rotates, for example, 30 ° is connected to the input port 35. On the other hand, the output port 36 is connected to the spark plug 6, the fuel injection valve 13, the throttle valve driving actuator 18, the variable compression ratio mechanism A, and the variable valve timing mechanism B through corresponding drive circuits 38.

  2 is an exploded perspective view of the variable compression ratio mechanism A shown in FIG. 1, and FIG. 3 is a side sectional view of the internal combustion engine schematically shown. Referring to FIG. 2, a plurality of protrusions 50 spaced from each other are formed below both side walls of the cylinder block 2, and cam insertion holes 51 each having a circular cross section are formed in each protrusion 50. Has been. On the other hand, a plurality of protrusions 52 are formed on the upper wall surface of the crankcase 1 so as to be fitted between the corresponding protrusions 50 spaced apart from each other. Cam insertion holes 53 each having a circular cross section are formed.

  As shown in FIG. 2, a pair of camshafts 54 and 55 are provided, and on each camshaft 54 and 55, a circular cam 56 is rotatably inserted into each cam insertion hole 51. It is fixed. These circular cams 56 are coaxial with the rotational axes of the camshafts 54 and 55. On the other hand, an eccentric shaft 57 arranged eccentrically with respect to the rotation axis of each camshaft 54, 55 extends between the circular cams 56 as shown by hatching in FIG. A cam 58 is eccentrically mounted for rotation. As shown in FIG. 2, the circular cams 58 are disposed between the circular cams 56, and the circular cams 58 are rotatably inserted into the corresponding cam insertion holes 53.

  When the circular cams 56 fixed on the camshafts 54 and 55 are rotated in opposite directions as indicated by solid arrows in FIG. 3A from the state shown in FIG. In order to move toward the lower center, the circular cam 58 rotates in the opposite direction to the circular cam 56 in the cam insertion hole 53 as shown by the broken arrow in FIG. 3A, and is shown in FIG. As described above, when the eccentric shaft 57 moves to the lower center, the center of the circular cam 58 moves below the eccentric shaft 57.

  3A and 3B, the relative positions of the crankcase 1 and the cylinder block 2 are determined by the distance between the center of the circular cam 56 and the center of the circular cam 58. The cylinder block 2 moves away from the crankcase 1 as the distance between the center and the center of the circular cam 58 increases. When the cylinder block 2 moves away from the crankcase 1, the volume of the combustion chamber 5 increases when the piston 4 is positioned at the compression top dead center. Therefore, by rotating the camshafts 54 and 55, the piston 4 is compressed at the top dead center. The volume of the combustion chamber 5 when it is located at can be changed.

  As shown in FIG. 2, in order to rotate the camshafts 54 and 55 in the opposite directions, a pair of worm gears 61 and 62 having opposite spiral directions are attached to the rotation shafts of the drive motor 59, respectively. Gears 63 and 64 that mesh with the worm gears 61 and 62 are fixed to the ends of the camshafts 54 and 55, respectively. In this embodiment, by driving the drive motor 59, the volume of the combustion chamber 5 when the piston 4 is located at the compression top dead center can be changed over a wide range. The variable compression ratio mechanism A shown in FIGS. 1 to 3 shows an example, and any type of variable compression ratio mechanism can be used.

  4 shows a variable valve timing mechanism B provided for the camshaft 70 for driving the intake valve 7 in FIG. As shown in FIG. 4, the variable valve timing mechanism B is attached to one end of the camshaft 70 to change the cam phase of the camshaft 70, and the camshaft 70 and intake valve 7 valve lifters. The cam operating angle changing portion B2 is disposed between the cam shafts 70 and changes the cam operating angle of the camshaft 70 to a different operating angle and transmits the cam operating angle to the intake valve 7. FIG. 4 shows a side sectional view and a plan view of the cam working angle changing portion B2.

  First, the cam phase changing portion B1 of the variable valve timing mechanism B will be described. The cam phase changing portion B1 is rotated by a crankshaft of the engine via a timing belt in a direction indicated by an arrow, a timing pulley 71, A cylindrical housing 72 that rotates together, a rotating shaft 73 that rotates together with the camshaft 70 and that can rotate relative to the cylindrical housing 72, and an outer peripheral surface of the rotating shaft 73 from an inner peripheral surface of the cylindrical housing 72 And a plurality of partition walls 74 extending between the partition walls 74 and the outer peripheral surface of the rotating shaft 73 to the inner peripheral surface of the cylindrical housing 72. An advance hydraulic chamber 76 and a retard hydraulic chamber 77 are formed respectively.

  The hydraulic oil supply control to the hydraulic chambers 76 and 77 is performed by a hydraulic oil supply control valve 78. The hydraulic oil supply control valve 78 includes hydraulic ports 79 and 80 connected to the hydraulic chambers 76 and 77, a hydraulic oil supply port 82 discharged from the hydraulic pump 81, a pair of drain ports 83 and 84, And a spool valve 85 for controlling communication between the ports 79, 80, 82, 83, and 84.

  When the cam phase of the camshaft 70 is to be advanced, the spool valve 85 is moved downward in FIG. 4, and the hydraulic oil supplied from the supply port 82 enters the advance angle hydraulic chamber 76 via the hydraulic port 79. The hydraulic oil in the retarding hydraulic chamber 77 is supplied and discharged from the drain port 84. At this time, the rotation shaft 73 is rotated relative to the cylindrical housing 72 in the arrow X direction.

  On the other hand, when the cam phase of the camshaft 70 is to be retarded, the spool valve 85 is moved upward in FIG. 4 and the hydraulic oil supplied from the supply port 82 is used for retarding via the hydraulic port 80. The hydraulic oil in the advance hydraulic chamber 76 is discharged from the drain port 83 while being supplied to the hydraulic chamber 77. At this time, the rotating shaft 73 is rotated relative to the cylindrical housing 72 in the direction opposite to the arrow X.

  If the spool valve 85 is returned to the neutral position shown in FIG. 4 while the rotation shaft 73 is rotated relative to the cylindrical housing 72, the relative rotation operation of the rotation shaft 73 is stopped, and the rotation shaft 73 is The relative rotation position at that time is held. Therefore, the cam phase change portion B1 can advance or retard the cam phase of the camshaft 70 by a desired amount. That is, the valve opening timing of the intake valve 7 can be arbitrarily advanced or retarded by the cam phase changing unit B1.

  Next, the cam working angle changing portion B2 of the variable valve timing mechanism B will be described. The cam working angle changing portion B2 is arranged in parallel with the camshaft 70 and is moved in the axial direction by the actuator 91; An intermediate cam 94 engaged with the cam 92 of the camshaft 70 and slidably fitted to an axially extending spline 93 formed on the control rod 90 and a valve lifter for driving the intake valve 7 And a swing cam 96 slidably fitted to a spirally extending spline 95 formed on the control rod 90. A cam 97 is formed on the swing cam 96. Has been.

  When the camshaft 90 is rotated, the intermediate cam 94 is always swung by a certain angle by the cam 92. At this time, the rocking cam 96 is also swung by a certain angle. On the other hand, the intermediate cam 94 and the swing cam 96 are supported so as not to move in the axial direction of the control rod 90. Therefore, when the control rod 90 is moved in the axial direction by the actuator 91, the swing cam 96 is intermediate. It is rotated relative to the cam 94.

  When the cam 97 of the swing cam 86 starts to engage with the valve lifter 24 when the cam 92 of the cam shaft 70 starts to engage with the intermediate cam 94 due to the relative rotational positional relationship between the intermediate cam 94 and the swing cam 96. As shown by a in FIG. 5B, the valve opening period and the lift of the intake valve 7 become the largest. On the other hand, when the swing cam 96 is rotated relative to the intermediate cam 94 in the direction of the arrow Y by the actuator 91, the cam 92 of the camshaft 70 is engaged with the intermediate cam 94 for a while. Thereafter, the cam 97 of the swing cam 96 is engaged with the valve lifter 24. In this case, as indicated by b in FIG. 5B, the valve opening period and the lift amount of the intake valve 7 are smaller than a.

  When the swing cam 96 is further rotated relative to the intermediate cam 94 in the direction of the arrow Y in FIG. 4, the valve opening period and the lift amount of the intake valve 7 are further reduced as indicated by c in FIG. . That is, the valve opening period of the intake valve 7 can be arbitrarily changed by changing the relative rotational position of the intermediate cam 94 and the swing cam 96 by the actuator 91. However, in this case, the lift amount of the intake valve 7 becomes smaller as the opening period of the intake valve 7 becomes shorter.

  Thus, the cam phase changing unit B1 can arbitrarily change the valve opening timing of the intake valve 7, and the cam operating angle changing unit B2 can arbitrarily change the valve opening period of the intake valve 7, so that the cam phase By both the change part B1 and the cam working angle change part B2, that is, by the variable valve timing mechanism B, the valve opening timing and the valve opening period of the intake valve 7, that is, the valve opening timing and the valve closing timing of the intake valve 7 are set. It can be changed arbitrarily.

  The variable valve timing mechanism B shown in FIGS. 1 and 4 shows an example, and various types of variable valve timing mechanisms other than the examples shown in FIGS. 1 and 4 can be used.

  Next, the meanings of terms used in the present application will be described with reference to FIG. 6 (A), (B), (C), and (D) show an engine having a combustion chamber volume of 50 ml and a piston stroke volume of 500 ml for the sake of explanation. In (A), (B), (C), and (D), the combustion chamber volume represents the volume of the combustion chamber when the piston is located at the compression top dead center.

  FIG. 6A explains the mechanical compression ratio. The mechanical compression ratio is a value mechanically determined only from the stroke volume of the piston and the combustion chamber volume during the compression stroke, and this mechanical compression ratio is expressed by (combustion chamber volume + stroke volume) / combustion chamber volume. In the example shown in FIG. 6A, this mechanical compression ratio is (50 ml + 500 ml) / 50 ml = 11.

  FIG. 6B describes the actual compression ratio. This actual compression ratio is a value determined from the actual piston stroke volume and the combustion chamber volume from when the compression action is actually started until the piston reaches top dead center, and this actual compression ratio is (combustion chamber volume + actual (Stroke volume) / combustion chamber volume. That is, as shown in FIG. 6B, even if the piston starts to rise in the compression stroke, the compression operation is not performed while the intake valve is open, and the actual compression is performed from the time when the intake valve is closed. The action begins. Therefore, the actual compression ratio is expressed as described above using the actual stroke volume. In the example shown in FIG. 6B, the actual compression ratio is (50 ml + 450 ml) / 50 ml = 10.

  FIG. 6C illustrates the expansion ratio. The expansion ratio is a value determined from the stroke volume of the piston and the combustion chamber volume during the expansion stroke, and this expansion ratio is expressed by (combustion chamber volume + stroke volume) / combustion chamber volume. In the example shown in FIG. 6C, this expansion ratio is (50 ml + 500 ml) / 50 ml = 11.

  FIG. 6D illustrates the actual expansion ratio. This actual expansion ratio is a value determined from the actual exhaust stroke volume and the combustion chamber volume from when the expansion action is started until the expansion action is actually finished, that is, until the exhaust valve 9 is opened. The ratio is expressed as (combustion chamber volume + actual exhaust stroke volume) / combustion chamber volume. In the example shown in FIG. 6D, the actual expansion ratio is (50 ml + 450 ml) / 50 ml = 10.

  Next, an ultra-high expansion ratio cycle used in the present invention will be described with reference to FIGS. FIG. 7 shows the relationship between the theoretical thermal efficiency and the actual expansion ratio, and FIG. 8 shows a comparison between a normal cycle and an ultrahigh expansion ratio cycle that are selectively used according to the load in the present invention.

  FIG. 8A shows a normal cycle when the intake valve closes near the bottom dead center and the compression action by the piston is started from the vicinity of the compression bottom dead center. In the example shown in FIG. 8 (A), the combustion chamber volume is set to 50 ml and the stroke volume of the piston is set to 500 ml, as in the examples shown in FIGS. 6 (A), (B), (C) and (D). ing. As can be seen from FIG. 8A, in a normal cycle, the mechanical compression ratio is (50 ml + 500 ml) / 50 ml = 11, the actual compression ratio is almost 11, and the expansion ratio and the actual expansion ratio are also (50 ml + 500 ml) / 50 ml = 11. It becomes. That is, in a normal internal combustion engine, the mechanical compression ratio, the actual compression ratio, the expansion ratio, and the actual expansion ratio are substantially equal.

  The solid line in FIG. 7 shows the change in the theoretical thermal efficiency when the actual compression ratio and the actual expansion ratio are substantially equal, that is, in a normal cycle. In this case, it can be seen that the theoretical thermal efficiency increases as the actual expansion ratio increases, that is, as the actual compression ratio increases. Therefore, in order to increase the theoretical thermal efficiency in a normal cycle, it is only necessary to increase the actual compression ratio. However, the actual compression ratio can only be increased to a maximum of about 12 due to the restriction of the occurrence of knocking at the time of engine high load operation, and thus the theoretical thermal efficiency cannot be sufficiently increased in a normal cycle.

  On the other hand, in this situation, the present inventor has studied to increase the theoretical thermal efficiency by strictly dividing the mechanical compression ratio and the actual compression ratio, and as a result, the theoretical thermal efficiency is governed by the actual expansion ratio, and the theoretical thermal efficiency On the other hand, it was found that the actual compression ratio has little influence. That is, if the actual compression ratio is increased, the explosive force is increased, but a large amount of energy is required for compression. Thus, even if the actual compression ratio is increased, the theoretical thermal efficiency is hardly increased.

  On the other hand, when the actual expansion ratio is increased, the period during which the pressing force acts on the piston during the expansion stroke becomes longer, and thus the period during which the piston applies the rotational force to the crankshaft becomes longer. Accordingly, the theoretical thermal efficiency increases as the actual expansion ratio increases. The broken line in FIG. 7 indicates the theoretical thermal efficiency when the actual expansion ratio is increased while the actual compression ratio is fixed at 10. In this way, the theoretical thermal efficiency increases when the actual expansion ratio is increased while the actual compression ratio is maintained at a low value, and the theory when the actual compression ratio is increased together with the actual expansion ratio as shown by the solid line in FIG. It can be seen that there is no significant difference from the increase in thermal efficiency.

  Thus, if the actual compression ratio is maintained at a low value, knocking does not occur. Therefore, if the actual expansion ratio is increased with the actual compression ratio maintained at a low value, the theoretical thermal efficiency is prevented while preventing knocking. Can be greatly increased. FIG. 8B shows an example of using the variable compression ratio mechanism A and the variable valve timing mechanism B to increase the actual expansion ratio while maintaining the actual compression ratio at a low value.

  Referring to FIG. 8B, in this example, the variable compression ratio mechanism A reduces the combustion chamber volume from 50 ml to 20 ml. On the other hand, the variable valve timing mechanism B delays the closing timing of the intake valve until the actual piston stroke volume is reduced from 500 ml to 200 ml. As a result, in this example, the actual compression ratio is (20 ml + 200 ml) / 20 ml = 11.

  On the other hand, FIG. 8B shows a case where the exhaust valve 9 opens near the bottom dead center and a case where the exhaust valve 9 opens when the stroke volume is 450 ml. When the exhaust valve 9 is opened near the bottom dead center, the actual expansion ratio is (20 ml + 500 ml) / 20 ml = 26, and when the exhaust valve 9 is opened when the stroke volume is 450 ml, the actual expansion ratio is ( 20 ml + 450 ml) / 20 ml = 23.5. In the normal cycle shown in FIG. 8A, the actual compression ratio is approximately 11 and the actual expansion ratio is 11 as described above. Compared to this case, the actual expansion ratio is greater in the case shown in FIG. It can be seen that only is increased to 26 or 23.5. This is why it is called an ultra-high expansion ratio cycle.

  Generally speaking, in an internal combustion engine, the lower the engine load, the worse the thermal efficiency. Therefore, in order to improve the thermal efficiency during engine operation, that is, to improve fuel efficiency, it is necessary to improve the thermal efficiency when the engine load is low. Necessary. On the other hand, in the ultra-high expansion ratio cycle shown in FIG. 8B, since the actual piston stroke volume during the compression stroke is reduced, the amount of intake air that can be sucked into the combustion chamber 5 is reduced. The expansion ratio cycle can only be adopted when the engine load is relatively low. Therefore, in the present invention, when the engine load is relatively low, the super high expansion ratio cycle shown in FIG. 8B is used, and during the high engine load operation, the normal cycle shown in FIG. 8A is used.

Next, general operation control when the three-way catalyst is sufficiently activated will be described with reference to FIG.
FIG. 9 shows each of the mechanical compression ratio, the actual expansion ratio, the closing timing of the intake valve 7, the actual compression ratio, the intake air amount, the opening degree of the throttle valve 19 and the pumping loss according to the engine load at a certain engine speed. Changes are shown. In the embodiment according to the present invention, the average air-fuel ratio in the normal combustion chamber 5 is the air-fuel ratio sensor 23 so that unburned HC, CO and NO _x in the exhaust gas can be simultaneously reduced by the three-way catalyst in the catalytic converter 22. Feedback control to the stoichiometric air-fuel ratio based on the output signal.

  As described above, the normal cycle shown in FIG. 8 (A) is executed during engine high load operation. Accordingly, as shown in FIG. 9, the actual expansion ratio is low because the mechanical compression ratio is lowered at this time, and as shown by the solid line in FIG. 9, the closing timing of the intake valve 7 is advanced as shown by the solid line in FIG. It has been. Further, at this time, the amount of intake air is large, and at this time, the opening degree of the throttle valve 19 is kept fully open or almost fully open, so that the pumping loss is zero.

  On the other hand, as shown by the solid line in FIG. 9, when the engine load becomes low, the closing timing of the intake valve 7 is delayed in order to reduce the intake air amount. Further, at this time, as shown in FIG. 9, the mechanical compression ratio is increased as the engine load is lowered so that the actual compression ratio is kept substantially constant. Therefore, the actual expansion ratio is also increased as the engine load is lowered. At this time as well, the throttle valve 19 is kept fully open or almost fully open, so the amount of intake air supplied into the combustion chamber 5 changes the closing timing of the intake valve 7 regardless of the throttle valve 19. Is controlled by that. Also at this time, the pumping loss becomes zero.

  As described above, when the engine load is reduced from the engine high load operation state, the mechanical compression ratio is increased as the intake air amount is decreased while the actual compression ratio is substantially constant. That is, the volume of the combustion chamber 5 when the piston 4 reaches the compression top dead center is decreased in proportion to the decrease in the intake air amount. Therefore, the volume of the combustion chamber 5 when the piston 4 reaches the compression top dead center changes in proportion to the intake air amount. At this time, since the air-fuel ratio in the combustion chamber 5 is the stoichiometric air-fuel ratio, the volume of the combustion chamber 5 when the piston 4 reaches the compression top dead center changes in proportion to the fuel amount. Become.

When the engine load is further reduced, the mechanical compression ratio is further increased, and when the engine load is lowered to the medium load L ₁ slightly close to the low load, the mechanical compression ratio reaches a limit mechanical compression ratio that is a structural limit of the combustion chamber 5. When the mechanical compression ratio reaches the limit mechanical compression ratio, the mechanical compression ratio is held at the limit mechanical compression ratio in a region where the load is lower than the engine load L ₁ when the mechanical compression ratio reaches the limit mechanical compression ratio. Therefore, the mechanical compression ratio is maximized and the actual expansion ratio is maximized at the time of low engine load operation and low engine load operation, that is, at the engine low load operation side. In other words, the mechanical compression ratio is maximized so that the maximum actual expansion ratio is obtained on the engine low load operation side.

On the other hand, in the embodiment shown in FIG. 9, even when the engine load becomes lower than L ₁ , as shown by the solid line in FIG. 9, the closing timing of the intake valve 7 is delayed as the engine load becomes lower. L closing timing of the intake valve 7 and decreases to ₂ becomes the limit closing timing enabling control of the amount of intake air fed into the combustion chamber 5. When the closing timing of the intake valve 7 reaches the limit closing timing, the closing timing of the intake valve 7 in a region where the load is lower than the engine load L ₂ when the closing timing of the intake valve 7 reaches the closing timing. Is held at the limit valve closing timing.

When the closing timing of the intake valve 7 is held at the limit closing timing, the amount of intake air can no longer be controlled by changing the closing timing of the intake valve 7. In the embodiment shown in FIG. 9, at this time, that is, in a region where the load is lower than the engine load L ₂ when the closing timing of the intake valve 7 reaches the limit closing timing, the throttle valve 19 supplies the fuel into the combustion chamber 5. The amount of intake air to be controlled is controlled. However, when the intake air amount is controlled by the throttle valve 19, the pumping loss increases as shown in FIG.

On the other hand, the engine load as shown in Figure 9 is the actual compression ratio at high engine high load operation side of the L ₁ is maintained in the actual compression ratio substantially the same for the same engine speed. On the other hand, when the engine load is lower than L ₂ , that is, when the mechanical compression ratio is maintained at the limit mechanical compression ratio, the actual compression ratio is determined by the closing timing of the intake valve 7, and the engine load is L ₁ and L _When the valve closing timing of the intake valve 7 is delayed as in 2, the actual compression ratio decreases, and the valve closing timing of the intake valve 7 becomes the limit valve closing as in the operation region where the engine load is lower than L _2. If it is held at the time, the actual compression ratio is kept constant.

On the other hand, as shown by the broken line in FIG. 9, the intake air amount can be controlled without depending on the throttle valve 19 by advancing the closing timing of the intake valve 7 as the engine load becomes lower. Accordingly, when expressing the case shown in FIG. 9 so as to include both the case indicated by the solid line and the case indicated by the broken line, in the embodiment according to the present invention, the valve closing timing of the intake valve 7 becomes smaller as the engine load becomes lower. so that is moved in a direction away from intake bottom dead center BDC until the limit closing timing L ₂ enabling control of the amount of intake air fed into the combustion chamber.

  Now, when the temperature of the three-way catalyst in the catalytic converter 22 falls to below the activation temperature, the exhaust gas is not purified, and therefore the three-way catalyst needs to be kept at the activation temperature or higher. On the other hand, as can be seen from FIG. 7, when the actual expansion ratio decreases, the theoretical thermal efficiency decreases and the exhaust gas temperature increases. Therefore, in the present invention, when the temperature of the three-way catalyst is predicted to fall below the activation temperature, the actual expansion ratio is lowered, thereby increasing the exhaust gas temperature to keep the temperature of the three-way catalyst above the activation temperature. ing.

  By the way, if the actual compression ratio is lowered when the actual expansion ratio is lowered, the ignition and combustion of the fuel will be worsened. Therefore, in the present invention, at this time, the actual expansion ratio is decreased while maintaining the same actual compression ratio or increasing the actual compression ratio.

  FIG. 10 shows an embodiment in which the actual expansion ratio is lowered by lowering the mechanical compression ratio when the temperature of the three-way catalyst is predicted to fall below the activation temperature. In FIG. 10, the solid line indicates the solid line in FIG. 9, that is, each value when the three-way catalyst is activated, and the broken line in FIG. 10 indicates each value when the temperature of the three-way catalyst is raised. Is shown.

  As can be seen from FIG. 10, in this embodiment, when the temperature of the three-way catalyst is to be increased, the mechanical compression ratio is decreased from the value shown by the solid line to the value shown by the broken line, and at this time, the actual expansion ratio is changed from the value shown by the solid line to the broken line. Reduced to the indicated value. On the other hand, in this embodiment, at this time, the actual compression ratio is increased from the value shown by the solid line to the value shown by the broken line. The opening of the throttle valve 19 is reduced from the solid line to the broken line.

  FIG. 11 shows another embodiment in which the expansion ratio is lowered by lowering the mechanical compression ratio when the temperature of the three-way catalyst is predicted to fall below the activation temperature. In FIG. 11, the solid line indicates the solid line in FIG. 9, that is, each value when the three-way catalyst is activated, and the broken line in FIG. 11 indicates each value when the temperature of the three-way catalyst is raised. Is shown.

  Also in this embodiment, when the temperature of the three-way catalyst is to be increased, the mechanical compression ratio is reduced from the value indicated by the solid line to the value indicated by the broken line, and at this time, the actual expansion ratio is reduced from the value indicated by the solid line to the value indicated by the broken line. . Also in this embodiment, at this time, the actual compression ratio is increased from the value shown by the solid line to the value shown by the broken line. The opening of the throttle valve 19 is reduced from the solid line to the broken line.

In this embodiment, unlike the embodiment shown in FIG. 10, the mechanical compression ratio is lowered to raise the temperature of the three-way catalyst only when the engine load is lower than a predetermined load L _0. The lower the is, the more the amount of decrease in the mechanical compression ratio is increased. That is, when the engine load is higher than L _0, the temperature of the three-way catalyst is considered not to be lower than the activation temperature. Therefore, in this embodiment, when the engine load is higher than L ₀ , it depends on the temperature of the three-way catalyst. Therefore, the temperature raising action of the three-way catalyst is not performed.

On the other hand, in a region where the engine load is lower than L ₀ , when the engine load is reduced, the exhaust gas temperature is lowered and the exhaust gas amount is reduced. Therefore, in order to keep the temperature of the three-way catalyst at the activation temperature when the temperature of the three-way catalyst is predicted to fall below the activation temperature, it is necessary to raise the exhaust gas temperature as the engine load decreases. Therefore, in this embodiment, when the temperature of the three-way catalyst is predicted to fall below the activation temperature, the amount of decrease in the mechanical compression ratio is increased as the engine load is reduced, and as a result, the amount of decrease in the actual expansion ratio is reduced as the engine load is reduced. Is increased.

FIG. 12 shows an operation control routine applicable to any of the embodiments shown in FIGS. Referring to FIG. 12, first, at step 100, the temperature TC of the three-way catalyst is estimated from the output signal of the temperature sensor 24. Next, at step 101, it is determined whether or not the temperature T _{0 of} the three-way catalyst is predicted to drop below the activation temperature, for example, a temperature T ₀ or less that is slightly higher than the temperature at which the three-way catalyst is activated. The When TC ≧ T ₀ , that is, when the three-way catalyst is sufficiently activated, the routine proceeds to step 102 where the operation control shown in FIG. 9 is performed.

  That is, at step 102, the target actual compression ratio CP is calculated. Next, at step 103, the closing timing IC of the intake valve 7 is calculated from the map shown in FIG. That is, the closing timing IC of the intake valve 7 necessary for supplying the required intake air amount into the combustion chamber 5 as a function of the engine load L and the engine speed N is in the form of a map as shown in FIG. The valve closing timing IC of the intake valve 7 is calculated from this map.

  Next, at step 104, the mechanical compression ratio CR is calculated. Next, at step 105, the opening of the throttle valve 17 is calculated. The opening θ of the throttle valve 17 is stored in advance in the ROM 32 as a function of the engine load L and the engine speed N in the form of a map as shown in FIG. Next, at step 110, the variable compression ratio mechanism A is controlled so that the mechanical compression ratio becomes the mechanical compression ratio CR, and the variable valve timing mechanism B is controlled so that the closing timing of the intake valve 7 becomes the closing timing IC, The throttle valve 17 is controlled so that the opening degree of the throttle valve 17 becomes the opening degree θ.

On the other hand, when it is determined in step 101 that TC <T ₀ , that is, when it is predicted that the temperature of the three-way catalyst will fall below the activation temperature, the routine proceeds to step 106 and the operation control indicated by the broken line in FIG. Switched. In the embodiment shown in FIG. 11, when TC <T ₀ when the engine load is lower than L ₀ , the routine proceeds to step 106 and is switched to the operation control indicated by the broken line in FIG.

  That is, first, at step 106, the target actual compression ratio CP 'is calculated. Next, at step 107, the closing timing IC 'of the intake valve 7 is calculated from the map shown in FIG. That is, also in this case, the closing timing IC ′ of the intake valve 7 necessary for supplying the required intake air amount into the combustion chamber 5 is shown as a function of the engine load L and the engine speed N in FIG. The map is stored in advance in the ROM 32, and the closing timing IC ′ of the intake valve 7 is calculated from this map.

  Next, at step 108, the mechanical compression ratio CR 'is calculated. Next, at step 109, the opening of the throttle valve 19 is calculated. The opening θ ′ of the throttle valve 19 is also stored in advance in the ROM 32 as a function of the engine load L and the engine speed N in the form of a map as shown in FIG. Next, at step 110, the variable compression ratio mechanism A is controlled so that the mechanical compression ratio becomes the mechanical compression ratio CR ′, and the variable valve timing mechanism B is controlled so that the closing timing of the intake valve 7 becomes the closing timing IC ′. Then, the throttle valve 19 is controlled so that the opening degree of the throttle valve 19 becomes the opening degree θ ′. At this time, the ignition timing can be retarded in order to further increase the exhaust gas temperature.

  15 to 17 show still another embodiment. In this embodiment, as shown in FIG. 15, a variable valve timing mechanism B ′ having the same structure as the variable valve timing mechanism B provided for the intake valve 7 is also provided for the exhaust valve 9. Therefore, this variable valve timing mechanism B 'can control the closing timing of the exhaust valve 9 and can individually control the opening timing of the exhaust valve 9.

In this embodiment, when it is predicted that the temperature of the three-way catalyst will fall below the activation temperature, the opening timing EO of the exhaust valve 9 is set to the normal EO ₀ as shown in FIG. 16 without reducing the mechanical compression ratio. From EO ₁ to EO ₁ , thereby reducing the actual expansion ratio.

FIG. 17 shows an operation control routine. Referring to FIG. 17, first, at step 200, the temperature TC of the three-way catalyst is estimated from the output signal of the temperature sensor 24. Next, at step 201, it is judged if the temperature TC of the three-way catalyst has fallen below the temperature T ₀ that is predicted to drop below the activation temperature. When TC ≧ T ₀ , that is, when the three-way catalyst is sufficiently activated, the routine proceeds to step 202 where the valve opening timing EO of the exhaust valve 9 is made the normal valve opening timing EO ₀ shown in FIG. Next, the routine proceeds to step 204.

On the other hand, when it is determined in step 201 that TC <T ₀ , that is, when it is predicted that the temperature of the three-way catalyst will fall below the activation temperature, the routine proceeds to step 203 where the opening timing EO of the exhaust valve 9 is set. It is advanced to EO ₁ shown in FIG. At this time, the ignition timing can be retarded to further increase the exhaust gas temperature. Next, the routine proceeds to step 204.

After step 204, the operation control shown in FIG. 9 is performed. That is, in step 204, the target actual compression ratio CP is calculated. Next, at step 205, the closing timing IC of the intake valve 7 is calculated from the map shown in FIG. Next, at step 206, the mechanical compression ratio CR is calculated. Next, at step 207, the opening degree θ of the throttle valve 19 is calculated from the map shown in FIG. Next, at step 208, the variable compression ratio mechanism A is controlled so that the mechanical compression ratio becomes the mechanical compression ratio CR, and the variable valve timing mechanism B is controlled so that the closing timing of the intake valve 7 becomes the closing timing IC, The variable valve timing mechanism B is controlled so that the valve opening timing EO of the exhaust valve 9 becomes the valve opening timing EO ₀ or EO _1, and the throttle valve 19 is controlled so that the opening degree of the throttle valve 19 becomes the opening degree θ. The

  As described above, in the ultra-high expansion ratio cycle shown in FIG. 8B, the actual expansion ratio is set to 26 or 23.5. This actual expansion ratio is preferably as high as possible, but if it is 20 or more, a considerably high theoretical thermal efficiency can be obtained. Therefore, in the present invention, the variable compression ratio mechanism A is formed so that the actual expansion ratio is 20 or more.

1 is an overall view of a spark ignition internal combustion engine. It is a disassembled perspective view of a variable compression ratio mechanism. 1 is a schematic side sectional view of an internal combustion engine. It is a figure which shows a variable valve timing mechanism. It is a figure which shows the lift amount of an intake valve and an exhaust valve. It is a figure for demonstrating a mechanical compression ratio, an actual compression ratio, an expansion ratio, and an actual expansion ratio. It is a figure which shows the relationship between theoretical thermal efficiency and an actual expansion ratio. It is a figure for demonstrating a normal cycle and a super-high expansion ratio cycle. It is a figure which shows changes, such as a mechanical compression ratio according to an engine load. It is a figure which shows changes, such as a mechanical compression ratio according to an engine load. It is a figure which shows changes, such as a mechanical compression ratio according to an engine load. It is a flowchart for performing operation control. It is a figure which shows maps, such as valve closing timing IC of an intake valve. It is a figure which shows maps, such as valve closing timing IC 'of an intake valve. It is a general view which shows another Example of a spark ignition type internal combustion engine. It is a figure which shows the valve opening time EO of an exhaust valve. It is a flowchart for performing operation control.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 Crankcase 2 Cylinder block 3 Cylinder head 4 Piston 5 Combustion chamber 7 Intake valve A Variable compression ratio mechanism B Variable valve timing mechanism

Claims (3)

A variable compression ratio mechanism capable of changing the mechanical compression ratio, a variable valve timing mechanism capable of controlling the closing timing of the intake valve, a throttle valve disposed in the engine intake passage, and a catalyst disposed in the engine exhaust passage When the catalyst is activated, the mechanical compression ratio is increased to the maximum mechanical compression ratio as the engine load is reduced, and the intake valve closing timing is the intake bottom dead center as the engine load is reduced. The spark ignition type internal combustion engine moved in a direction away from the engine is provided with a predicting means for predicting the temperature of the catalyst arranged in the engine exhaust passage, and the temperature of the catalyst is predicted to be lowered below the activation temperature. time in accordance with the engine load in order to reduce the actual expansion ratio while increasing the while being or actual compression ratio holding the actual compression ratio at the same decreases, the amount of decrease in the mechanical compression ratio is increased Amount to shift the closing timing of the intake valve in a direction approaching the intake bottom dead center is large, spark ignition type internal combustion engine the amount of decrease in the opening degree of the throttle valve is increased. The spark ignition internal combustion engine according to claim 1, wherein the ignition timing is retarded when the temperature of the catalyst is predicted to fall below the activation temperature . 2. The spark ignition internal combustion engine according to claim 1 , wherein an expansion ratio at the maximum mechanical compression ratio is 20 or more .

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